Free Piston Engine Generator

ABSTRACT

A free-piston engine generator comprising an engine cylinder, a piston configured to move within the cylinder, a cylinder housing having a bore for receiving the engine cylinder and a plurality of magnetisable elements arranged within the cylinder housing to be adjacent the cylinder along at least a portion of its length.

The present invention relates to a free piston engine generator and amethod of manufacturing an engine generator. In particular, the presentinvention relates to a free piston engine generator which has aconstruction arranged to optimise the efficiency of the generator and toprovide good control of piston position and piston motion by thegenerator.

Electrical power can be generated using a linear generator coupled to afree piston engine, wherein the linear movement of the reciprocatingpiston through one or more electrical coils generates magnetic fluxchange, as disclosed in U.S. Pat. No. 7,318,506, for example. As thepiston moves within the cylinder past the coils it interacts with aswitched magnetic flux within the stator elements to generate electricalpower that can be used for useful work or stored for later use.

However, the efficiency of such an electrical power generation system ishighly dependent on the thickness of the cylinder wall and the proximityof the power generation electrical elements to the piston reciprocatingwithin the engine cylinder.

According to the present invention there is provided a free-pistonengine generator comprising an engine cylinder, a piston configured tomove within the cylinder, a cylinder housing having a bore for receivingthe engine cylinder and a plurality of magnetisable elements arrangedwithin the cylinder housing to be adjacent the cylinder along at least aportion of its length.

In the present invention, the particular arrangement of the cylinderhousing which allows the plurality of magnetisable elements to bepositioned adjacent the cylinder optimises the efficiency of the freepiston engine generator and provides good control of the piston positionand piston motion by the generator.

Preferably, the cylinder housing has one or more recesses that permitthe plurality of magnetisable elements to be positioned adjacent thecylinder. Preferably, the magnetisable elements are positioned in directphysical contact with the cylinder. Preferably, one or more of themagnetisable elements acts as a stator in the generator.

Preferably, the cylinder is secured within the cylinder housing by anadhesive which provides thermal insulation between the cylinder andcylinder housing.

Preferably, the cylinder housing comprises a plurality of elements thatare assembled coaxially onto the cylinder.

Preferably, the cylinder housing includes cooling means for cooling thecylinder.

Preferably, the cylinder has a wall thickness that is less than 5% ofthe cylinder's internal diameter, wherein the wall thickness istypically less than about 2 mm.

Preferably, the piston comprises alternating laminated magnetisable coreelements and non-magnetising spacer elements. Preferably, the intakemeans are located at a central position along the cylinder, whichsimplifies the engine arrangement and makes this arrangement morecompact by allowing a common intake means to supply fluid into eachcombustion chamber.

Furthermore, by positioning the intake means at a position removed fromthe exhaust valve, scavenging of burned gases can be greatly improved bythe unidirectional scavenging flow provided within each combustionchamber, which in turns results in improved efficiency, reduced unburnedfuel hydrocarbons and lower costs.

Preferably, the intake means comprises both an air intake means and afuel injection means, so that fuel injection into a combustion chambermay occur during the admission of intake charge air. The intake meansmay also comprise a plurality of air intake means and at least one fuelinjection means. Providing the air intake means and fuel injection meanstogether in the intake means allows both these features to share acommon sliding port valve, each being recessed within the void behindthis sliding port valve. This results in a simpler and hence cheaperconstruction.

Preferably, the air intake means comprises a sliding port valve and asecondary valve such as a solenoid poppet valve, barrel valve or othervalve means arranged in series with the sliding port valve. Thesecondary valve can allow air into the chamber at any time when thesliding port valve is uncovered by the piston, which allows good controlof the expansion ratio in response to a combustion event, independentlyof the position of the piston within the limits defined by the openingand closing positions of the sliding port valve.

Preferably, the fuel injection means comprises two injectors arrangedone on each side of the air intake secondary valve to allow fuel to beinjected directly into the respective chamber independently of whetherthe intake secondary valve is open or closed. The injectors are,ideally, piezo-injectors, which allow for precise, low cost electronicactuation and control of the fuel injection.

Preferably, the fuel injection means is configured to inject fuelimmediately prior to the closing of the slide valve to ensure that fuelinjected cannot be carried to and out of the exhaust port by scavengingair intake charge before the exhaust valve is closed, reducinghydrocarbon emissions.

Preferably, ignition means are provided in each chamber to initiatecombustion of the compressed air-fuel mixture. Use of spark ignitionfuels and their related operating cycles inherently generate lessparticulate emissions than compression ignition fuels and cycles.

Preferably, an exhaust means is provided in each combustion chamber toallow for burnt gases to be exhausted from the chamber followingcombustion.

Preferably, the exhaust means is a solenoid poppet valve provided ineach combustion chamber, with the valves being coaxial with the cylindersuch that the limiting area in the exhaust flow may approach 40% of thecylinder bore section area, reducing exhaust gas back-pressure duringexhaust and scavenging.

Preferably, the cylinder has a length at least ten times greater thanits diameter, which provides reduced variability of compression ratio ineach cycle, resulting from a low rate of change of compression ratiowith piston displacement error at top dead centre.

Preferably, the piston is configured to be elongate and the enginecylinder has a bore dimensioned such that a compression ratio of between10:1 and 16:1 can be achieved. This is higher than can be achieved in aconventional spark ignition engine due to detonation (knocking).Preferably, the engine is a ‘flex-fuel’ engine operating on any mixtureof gasoline, anhydrous ethanol and hydrous ethanol. The compressionratio may be optimised by the engine management system according to theparticular ethanol/gasoline/water blend that is used.

Also, an expansion ratio greater than twice the compression ratio isobtained. A long expansion stroke allows more of the combustion energyto be transferred into the piston, and in addition allows more time forcontrol (i.e. to react to measured piston speed variability).

Preferably, the intake means is positioned a suitable distance from theexhaust valve to ensure that a compression ratio of between 10:1 and16:1 can be achieved.

According to the present invention there is also provided a method ofmanufacturing an engine generator, comprising providing a cylinderconfigured to accommodate at least one piston that is free toreciprocate within the cylinder, securing the cylinder within a cylinderhousing that has a bore arranged to receive and provide structuralsupport for the cylinder and arranging a plurality of magnetisableelements adjacent the cylinder such that, when the piston moves withinthe cylinder, it induces magnetic flux as it passes the plurality ofmagnetisable elements.

Preferably, one or more recesses in the cylinder housing are providedfor receiving the plurality of magnetisable elements.

Preferably, one or more sections of the cylinder housing along thelength of the cylinder are removed to expose one or more sections of thecylinder wall and the plurality of magnetisable elements are arrangedwithin the recesses such that they are in direct physical contact withthe cylinder wall.

According to the present invention there is also provided a method ofmanufacturing an engine, comprising providing a cylinder configured toaccommodate at least one piston that is free to reciprocate within thecylinder, extruding a cylinder housing that is arranged to retain andprovide structural support for the cylinder, and securing the cylinderwithin the cylinder housing such that the cylinder wall is reinforced bythe structure of the cylinder housing.

Preferably, the plurality of magnetisable elements are arranged toprovide load-bearing support to the cylinder. Preferably, themagnetisable elements are arranged to provide a force against thecylinder wall, for example they may be biased against the cylinder wall,or may be pre-loaded such that they apply a force against the cylinderwall when positioned adjacent it.

Preferably, the cylinder is secured within the cylinder housing anadhesive material on the outside of the cylinder, wherein the adhesivematerial provides thermal insulation between the cylinder and cylinderhousing.

Preferably, the cylinder housing is provided with cooling means forcooling the cylinder.

Preferably, the interior wall of the cylinder is coated with afriction-reducing material to reduce friction between the interior walland a piston passing along it.

Preferably, the thickness of the cylinder wall is less than 5% of thecylinder's internal diameter, wherein the thickness of the cylinder wallis typically less than about 2 mm.

According to the present invention there is also provided a vehiclehaving a free piston engine generator, as described above.

The construction of the engine generator of the present inventionprovides a number of important advantages over the typical free pistonengine generator construction, in which two cylinder heads are affixedto two cylinders, with a separate electrical machine assembly locatedaxially between the cylinders.

The cylinder housing is, preferably, formed by extrusion of a ductilematerial, such as aluminium alloy. Advantageously, the elongatedconstruction permits the use of extrusion manufacturing technology toform the cylinder housing, rather than casting or extensive CNCmachining technology used for conventional engines. Extrusion offers afaster manufacturing cycle time and higher tolerances before machiningoperations than casting, reducing finished part cost. Similarly, thecylinder may be formed from extrusion or other mature, low costtube-forming manufacturing technologies. This construction thereforereduces the overall cost of the engine's cylinder assembly.

Furthermore, the contiguous form of the cylinder housing, which remainsunbroken across the mid-section of the engine, ensures that bothcombustion chambers are coaxially aligned with high precision andprovides a continuous bearing surface for the piston to travel across.This permits the piston to move over and past a centrally disposedintake, as described herein, whilst minimising the amount of wear to theinner surface of the cylinder during the operating life of the enginegenerator.

Although the cylinder housing is, preferably, a single extruded element,it could, alternatively, be formed by the coaxial assembly of a stack ofdissimilar extruded elements onto a common cylinder. For example, twoextrusions may be placed either side of an intake means, wherein theextrusions are assembled coaxially onto the cylinder.

Ideally, the wall of the cylinder housing extrusion should besufficiently thick and/or strong that it is load-bearing to allow a muchthinner cylinder wall to provide wear and sealing surfaces than wouldotherwise be required. The cylinder housing, ideally, has sections ofmaterial removed along the length of the cylinder to form one or morerecesses that, ideally, expose the wall of the cylinder housed within.The recesses are formed through the cylinder housing, preferablyextending from the outer surface inwards, such that the recesses openoutwards. A plurality of magnetisable elements can be positioned inclose proximity to the cylinder by arranging them in the one or morerecesses, each magnetisable element preferably fixed directly to thewall of the cylinder, which separates them from the moving magneticcircuit elements of the piston.

The cylinder wall thickness dimension is an important determinant of theefficiency of the electrical machine, and should be as small as possiblefor high efficiency. By providing adequate load bearing strength usingthe cylinder housing and magnetisable stator elements, the cylinder wallis not required to bear cylinder fluid pressures and may be madeconsiderably thinner subject to manufacturing, assembly and wearconstraints.

The inner and outer surfaces of the cylinder provide substrates for wearand thermal coatings respectively. A thermal coating can be applied tothe cylinder outer surface in the form of an adhesive material toprovide a secure, insulating and load bearing bond between the cylinderand cylinder housing. Furthermore, the arrangement of securing acylinder within the cylinder housing provides the advantage that themating surfaces of the respective components do not need to be finishedto any particular standard, other than to allow the cylinder to befitted within the cylinder housing.

A free piston engine generator according to the present invention has anumber of applications. For example, it may be integrated in aseries-hybrid electric vehicle power train incorporating a transientelectrical power store and one or more drive motors suitable for use asan automotive power source in small passenger vehicles, whereinelectrical power generated by the free piston engine is accumulated inan electrical energy storage device on board the vehicle to be deliveredto the vehicle drive motors on demand.

As a power source for a small passenger vehicle, the present inventionpreferably runs on a two-stroke engine cycle with spark ignition, withfour cylinders being arranged in a planar configuration such that theengine might be transverse mounted beneath the front or rear seats ofthe vehicle, offering significantly more design flexibility to thelayout of the passenger and storage spaces compared to a conventionalinternal combustion engine.

Each cylinder includes a free piston whose movement induces electricalpower in a linear generator arranged around each cylinder, and whosemovement is controllable by various means including the timing of valveand ignition events, and by modulation of the power drawn from orsupplied to the piston on each stroke. The movement of pistons issynchronised such that the engine is fully balanced, wherein the piston,ideally, comprises alternating magnetisable elements and non-magnetisingspacer elements.

Furthermore, each cylinder is charged by means of an intake mechanismthat introduces fluid into the cylinder at a position distal from eachend of the cylinder. The intake mechanism includes a poppet valve andsliding port valve in series such that the timing of the intake flowevents may be controlled independently of the piston positions relativeto the cylinders. Exhaust gas leaves the cylinders from exhaust valvemechanisms located at the end of each cylinder.

The geometry of the cylinder and disposition of the intake and exhaustmechanisms are such that the exhaust scavenging is completed withlimited mixing between intake fluid and exhaust fluid. The combustionchamber geometry offers a low surface area-to-volume ratio, and lowconductivity materials are used in the piston crown and cylinder head,so that minimal heat is rejected from the engine. The cylinder andpiston geometry provides an expansion ratio which is at least two timesthe compression ratio.

The arrangement, and number, of cylinders used is, however, dependent onthe application and the engine operating cycle can also be varied fordifferent applications, for example: spark ignition internal combustion;homogeneous charge compression ignition internal combustion; andheterogeneous charge compression ignition. Some of the features of thepresent invention may also be embodied with an external combustioncycle. Examples of external combustion cycle embodiments include use ofthe present invention as a gas expander for fluid from a gas turbineexhaust, an organic rankine cycle or a Stirling cycle. In a Stirlingengine, heat from an external combustion source is supplied to thechamber containing compressed working fluid at top dead centre. Afterexpansion, the exhaust gases are expelled to a closed cooling chamberbefore being readmitted to the chamber through the intake means in aclosed circuit.

The fuel in various alternative embodiments may be hydrous ethanol,anhydrous ethanol-gasoline blends, or gasoline. The invention may alsobe embodied as using diesel, bio-diesel, methane (CNG, LNG or biogas) orother gaseous or liquid fuels. In an external combustion embodiment awide range of combustible fuels may be used.

Accordingly, in conjunction with an energy storage system to providepeak transient power output requirements, the present invention providesa low-cost, high efficiency power supply for small passenger vehicleautomotive applications, and many other applications where low cost andhigh efficiency are key design considerations, for example as a staticpower generator for distributed power generation.

An example of the present invention will now be described, withreference to the accompanying figures, in which:

FIG. 1 shows a longitudinal section through a cylinder having a pistonaccording to an example of the present invention;

FIG. 2 is a longitudinal section through the piston, showing theconstruction from planar elements;

FIG. 3 is a perpendicular section through the piston, showing theconcentric arrangement of the shaft and planar elements;

FIG. 4 is a sectional view of the cylinder of FIG. 3 illustrating themagnetic flux in switched stator elements caused by movement of thepiston according tot the present invention;

FIG. 5 a is a perpendicular section through a cylinder showing thelinear generator stator and the magnetic circuit formed by a permeableelement in the first piston;

FIG. 5 b is a perpendicular section of an alternative linear generatorstator arrangement for two adjacent cylinders wherein the lineargenerator stator and the magnetic circuit are formed by a permeableelement in the first piston;

FIG. 6 is a partial sectional view of the cylinder illustrating itsconstruction;

FIG. 7 is a more detailed longitudinal section of the intake poppetvalve, intake port valve and fuel injector arrangement during the intakecharge displacement scavenging phase;

FIG. 8 is a more detailed longitudinal section of the exhaust meansincluding the exhaust poppet valve and actuator during the exhaustphase;

FIG. 9 is a time-displacement plot showing the changing piston positionwithin a cylinder during a complete engine cycle, and the timing ofengine cycle events during this period;

FIG. 9 a is a table showing different compression ratio control meansthat may be employed to control the compression ratio in a typicalengine cycle;

FIG. 9 b is a flow chart corresponding to the table in FIG. 9 a;

FIG. 10 is a pressure-volume plot showing a typical cylinder pressureplot during a complete engine cycle;

FIG. 11 is a schematic longitudinal section through a cylinder at topdead centre, at the end of the compression phase and around the time ofspark ignition and initiation of the combustion event in the firstchamber;

FIG. 12 is a schematic longitudinal section through a cylinder mid waythrough the expansion phase of the first chamber;

FIG. 13 is a schematic longitudinal section through a cylinder at theend of the expansion phase, but before the intake poppet valve hasopened;

FIG. 14 is a schematic longitudinal section through a cylinder followingthe opening of the intake poppet valve to charge chamber 1, allowingintake charge fluid pressure to equalise the lower cylinder pressure inthe first chamber;

FIG. 15 is a schematic longitudinal section through a cylinder followingthe opening of the exhaust poppet valve, and whilst the intake poppetvalve remains open, scavenging the first chamber;

FIG. 16 is a schematic longitudinal section through a cylinder duringfuel injection into the first chamber after the intake poppet valve hasclosed;

FIG. 17 is a schematic longitudinal section through a cylinder duringlubricant injection onto the piston outer surface;

FIG. 18 is a schematic longitudinal section through a cylinder whilstthe exhaust poppet valve is open, and after the intake poppet valve andsliding port valve have closed such that continuing expulsion of exhaustgases from the first chamber is achieved by piston displacement;

FIG. 19 is a schematic longitudinal section through a cylinder mid waythrough the compression phase in the first chamber;

FIGS. 20A and 20B shows a section of a cylinder housing of an enginegenerator according to the present invention both with (A) and without(B) the electrical machine attached;

FIG. 20C shows a perpendicular section of a cylinder housing of anengine generator through plane X-X indicated on FIG. 20A;

FIG. 21 is a schematic perpendicular section of a four cylinder engineconstruction through the intake means including the electrical chargecompressor;

FIG. 22 is a schematic perpendicular section of a four cylinder engineconstruction through the electrical generator means; and

FIG. 23 is a schematic perpendicular section of a four cylinder engineconstruction through the exhaust means.

FIG. 1 shows an example of an engine according to the present invention,comprising a hollow linear cylinder 1. A piston 2 is provided within thecylinder 1, the piston 2 having a constant diameter that is configuredto be slightly smaller than the inside diameter of the cylinder 1, butonly to the extent that the piston 2 is free to move along the length ofthe cylinder 1. The piston 2 is otherwise constrained in coaxialalignment with the cylinder 1, thereby effectively partitioning thecylinder 1 into a first combustion chamber 3 and a second combustionchamber 4, each chamber having a variable volume depending on theposition of the piston 2 within the cylinder 1. No part of the piston 2extends outside the cylinder 1. Using the first chamber 3 as an example,each of the chambers 3, 4 has a variable height 3 a and a fixed diameter3 b.

The cylinder 1 is, preferably, rotationally symmetric about its axis andis symmetrical about a central plane perpendicular to its axis. Althoughother geometric shapes could potentially be used to perform theinvention, for example having square or rectangular section pistons, thearrangement having circular section pistons is preferred. The cylinder 1has a series of apertures 1 a, 1 b provided along its length and distalfrom the ends, preferably in a central location. Through motion of thepiston 2, the apertures 1 a, 1 b form a sliding port intake valve 6 a,which is arranged to operate in conjunction with an air intake 6 bprovided around at least a portion of the cylinder 1, as is described indetail below.

The cylinder 1 preferably has a length at least ten times greater thanits diameter to provide reduced variability of compression ratio in eachcycle, as a result of a low rate of change of compression ratio withpiston displacement error at top dead centre.

FIG. 2 shows a piston 2 having an outer surface 2 a and comprising acentral shaft 2 c onto which are mounted a series of cylindricalelements. These cylindrical elements may include a piston crown 2 d ateach end of the central shaft 2 c, each piston crown 2 d preferablyconstructed from a temperature resistant and insulating material such asceramic. The piston crown end surface 2 b is, preferably, slightlyconcave, reducing the surface area-to-volume ratios of the first andsecond chambers 3, 4 at top dead centre and thereby reducing heatlosses. Of course, if the cylinder was of a different geometry then theconfiguration of these elements would be adapted accordingly.

The piston crown 2 d includes oil control features 2 e to control thedegree of lubrication wetting of the cylinder 1 during operation of theengine. These features comprise a groove and an oil control ring as arecommonly employed in conventional internal combustion engines.

Laminated core elements 2 f are also mounted on the piston shaft 2 c.Each core element 2 f is constructed from laminations of a magneticallypermeable material, such as iron ferrite, to reduce eddy current lossesduring operation of the engine.

Spacer elements 2 g are also mounted on the piston shaft 2 c. Eachspacer element 2 g ideally has low magnetic permeability and ispreferably constructed from a lightweight material such as aluminiumalloy and has a void 2 h formed within it to further reduce its weightand hence reduce mechanical forces exerted on the engine utilising it.The spacer elements 2 g are included to fix the relative position ofeach of the core elements 2 f and also act to limit the loss of“blow-by” gases flowing out of each chamber 3, 4 through the gap betweenthe piston wall and cylinder wall, whilst keeping the overall mass ofthe piston 2 assembly to a minimum.

Bearing elements 2 i are also mounted on the piston shaft 2 c, locatedat approximately 25% and 75% of the length of the piston 2 to reduce therisk of thermally-induced distortion of the axis of the piston 2 causingit to lock in the cylinder 1 or otherwise damage the cylinder 1. Eachbearing element 2 i features a weight-reduction void 2 j and has adiameter very slightly larger than the core elements 2 f and the spacerelements 2 g. The bearing elements 2 i also have a profiled outersurface 2 k for bearing the weight of the piston 2, and any other sideloads present, whilst keeping frictional losses and wear to a minimum.The bearing element 2 i are preferably constructed from a hard, wearresistant material such as ceramic or carbon and the profiled outersurface 2 k may be coated in a low friction material. Alternatively,bearing elements may incorporate roller bearing features as are commonlyused in sliding applications.

Similar to the piston crown, or perhaps instead of, the bearing element2 i may also include oil control features to control the degree oflubrication wetting of the cylinder 1 during operation of the engine.These features comprise a groove and an oil control ring as are commonlyemployed in conventional internal combustion engines.

The total length of the piston is, preferably, five times its diameterand is at least sufficient to completely close the sliding port valvesuch that at no time does the sliding port valve allow combustionchambers 3 and 4 to communicate.

FIG. 3 is a sectional view of the piston 2, showing the piston shaft 2 cpassing through a core element 2 f. The piston shaft ends 2 l aremechanically deformed or otherwise fixed to the piston crowns 2 d suchthat the elements 2 f, 2 g, 2 i that are mounted to the piston shaft 2 care securely retained under the action of tension maintained in thepiston shaft 2 c.

The alternating arrangement of core elements 2 f and spacers 2 gpositions the core laminations 2 f at the correct pitch for efficientoperation as, for example, part of a linear switched reluctancegenerator machine comprising the moving piston 2 and a linear generatormeans, for example a plurality of coils spaced along the length of thecylinder within which the piston reciprocates.

FIG. 4 shows an example of linear generator means 9 provided around theoutside of the cylinder 1, along at least a portion of its length, forfacilitating the transfer of energy between the piston 2 and electricaloutput means 9 e. The linear generator means 9 includes a number ofcoils 9 a and a number of stators 9 c, in the form of magnetisableelements, alternating along the length of the linear generator means 9.

The linear generator means 9 may be of a number of different electricalmachine types, for example a linear switched reluctance generator. Inthe arrangement shown, coils 9 a are switched by switching device 9 b soas to induce magnetic fields within the magnetisable stators 9 c and thepiston core laminations 2 e. The magnetisable stators 9 c may belaminated or constructed from a soft magnetic composite (SMC) material,for example. In each approach the stators are constructed fromelectrically conducting and magnetisable elements separated bynon-conducting material which reduces heat losses from magneticallyinduced eddy currents.

The transverse magnetic flux created in the magnetisable stators 9 c andpiston core laminations 2 f under the action of the switched coils 9 ais also indicated in FIG. 4. The linear generator means 9 functions as alinear switched reluctance device, or as a linear switched flux device.Power is generated at the electrical output means 9 e as the fluxcircuits, established in the magnetisable stators 9 c and induced in thepiston core laminations 2 f, are cut by the motion of the piston 2. Thispermits a highly efficient electrical generation means without the useof permanent magnets, which may demagnetise under the high temperatureconditions within an internal combustion engine, and which mightotherwise add significant cost to the engine due the use of costly rareearth metals.

Additionally, a control module 9 d may be employed, comprising severaldifferent control means, as described below. The different control meansare provided to achieve the desired rate of transfer of energy betweenthe piston 2 and electrical output means 9 e in order to deliver themaximum electrical output whilst satisfying the desired motioncharacteristics of the piston 2, including compression rate and ratio,expansion rate and ratio, and piston dwell time at top dead centre ofeach chamber 3, 4.

A valve control means may be used to control the intake valve 6 c andthe exhaust valve 7 b. By controlling the closure of the exhaust valve 7b, the valve control means is able to control the start of thecompression phase. In a similar way, the valve control means can also beused to control exhaust gas recirculation (EGR), intake charge andcompression ratio.

A compression ratio control means that is appropriate to the type ofelectrical machine may also be employed. For example, in the case of aswitched reluctance machine, compression ratio control is partiallyachieved by varying the phase, frequency and current applied to theswitched coils 9 a. This changes the rate at which induced transverseflux is cut by the motion of the piston 2, and therefore changes theforce that is applied to the piston 2. Accordingly, the coils 9 a may beused to control the kinetic energy of the piston 2, both at the point ofexhaust valve 7 b closure and during the subsequent deceleration of thepiston 2.

A spark ignition timing control means may then be employed to respond toany residual cycle-to-cycle variability in the compression ratio toensure that the adverse impact of this residual variability on engineemissions and efficiency are minimised, as follows. Generally, theexpected compression ratio at the end of each compression phase is thetarget compression ratio plus an error that is related to systemvariability, such as the combustion event that occurred in the oppositecombustion chamber 3, 4, and the control system characteristics. Thespark ignition timing control means may adjust the timing of the sparkignition event in response to the measured speed and acceleration of theapproaching piston 2 to optimize the combustion event for the expectedcompression ratio at the end of each compression phase.

The target compression ratio will normally be a constant depending onthe fuel 5 a that is used. However, a compression ratio error may bederived from a +/−20% variation of the combustion chamber height 3 a.Hence if the target compression ratio is 12:1, the actual compressionratio may be in the range 10:1 to 15:1. Advancement or retardation ofthe spark ignition event by the spark ignition timing control means willtherefore reduce the adverse emissions and efficiency impact of thiserror.

Additionally, a fuel injection control means may be employed to controlthe timing of the injection of fuel 5 a so that it is injected into acombustion chamber 3, 4 immediately prior to the sliding port valve 6 aclosing to reduce hydrocarbon emissions during scavenging.

Furthermore, a temperature control means may be provided, including oneor more temperature sensors positioned in proximity to the coils 9 a,electronic devices and other elements sensitive to high temperatures, tocontrol the flow of cooling air in the system via the compressor 6 e inresponse to detected temperature changes. The temperature control meansmay be in communication with the valve control means to limit enginepower output when sustained elevated temperature readings are detectedto avoid engine damage.

Further sensors that may be employed by the control module 9 dpreferably include an exhaust gas (Lambda) sensor and an air flow sensorto determine the amount of fuel 5 a to be injected into a chamberaccording to the quantity of air added, for a given fuel type.Accordingly, a fuel sensor may also be employed to determine the type offuel being used.

FIG. 5 a shows a perpendicular section through one of the magnetisablestator elements 9 c, showing the arrangement of coils 9 a and stators 9c relative to each other. An alternative embodiment is shown in FIG. 5b, in which a single stator 9 c and coil 9 a are used to induce magneticflux in two adjacent pistons 2. This configuration has a cost advantagecompared to that shown in FIG. 5 a due to the reduced number of coils 9a required.

FIG. 6 is a sectional view of the cylinder 1, which is preferablyconstructed from a material of low magnetic permeability, such as analuminium alloy. The inner surface 1 c of the cylinder 1 has a coating 1e of a hard, wear-resistant material such as nickel silicon-carbide,reaction bonded silicon nitride, chrome plating, or other metallic,ceramic or other chemical coating. On the outer surface 1 d, aninsulator coating 1 f such as zirconium oxide or other sufficientlythermally insulating ceramic is applied. It will be apparent to askilled person that the whole cylinder has an identical construction tothis sectional view of the part of the cylinder close to the cylinderend 1 g.

FIG. 7 shows the intake means 6 provided around the cylinder 1, theintake means 6 comprising apertures 6 a, which are a corresponding sizeand align with the apertures 1 a, 1 b provided in the cylinder 1, and anair intake 6 b. The apertures 6 a in the intake means 6 are connected bya channel 6 h in which an intake poppet valve 6 c is seated. The channel6 h is of minimal volume, either having a short length, small crosssectional area or a combination of both, to minimise uncontrolledexpansion losses within the channel 6 h during the expansion phase.

The intake poppet valve 6 c seals the channel 6 h from an intakemanifold 6 f provided adjacent to the cylinder 1 as part of the airintake 6 b. The intake poppet valve 6 c is operated by a poppet valveactuator 6 d, which may be an electrically operated solenoid means orother suitable electrical or mechanical means.

When the sliding port intake valve 6 a and the intake poppet valve 6 care both open with respect to one of the first or second chambers 3, 4,the intake manifold 6 f is in fluid communication with that chamber viathe channel 6 h. The intake means 6 is preferably provided with a recess6 g arranged to receive the intake poppet valve 6 c when fully open toensure that fluid can flow freely through the channel 6 h.

The air intake 6 b also includes an intake charge compressor 6 e whichmay be operated electrically, mechanically, or under the action ofpressure waves originating from the air intake 6 b. The intake chargecompressor 6 e can also be operated under the action of pressure wavesoriginating from an exhaust means 7 provided at each end of the cylinder1, as described below, or by a conventional exhaust turbocharger device.The intake charge compressor 6 e may be a positive displacement device,centrifugal device, axial flow device, pressure wave device, or anysuitable compression device. The intake charge compressor 6 e elevatespressure in the intake manifold 6 f such that when the air intake 6 b isopened, the pressure in the intake manifold 6 f is greater than thepressure in the chamber 3, 4 connected to the intake manifold 6 f,thereby permitting a flow of intake charge fluid.

Fuel injection means 5 are also provided within the intake means 6, suchas a solenoid injector or piezo-injector 5. Although a centrallypositioned single fuel injector 5 may be adequate, there is preferably afuel injector 5 provided either side of the intake poppet valve 6 c andarranged proximate to the extremities of the sliding port valves 6 a.The fuel injectors 5 are preferably recessed in the intake means 6 suchthat the piston 2 may pass over and past the sliding port intake valves6 a and air intake 6 b without obstruction. The fuel injectors 5 areconfigured to inject fuel into the respective chambers 3, 4 through eachof the sliding port intake valves 6 a

Lubrication means 10 are also provided preferably recessed within theintake means 6 and arranged such that the piston 2 may pass over andpast the intake means 6 without obstruction, whereby the piston may belubricated.

FIG. 8 shows the exhaust means 7 provided at each end of the cylinder 1.The exhaust means 7 comprises a cylinder head 7 a removably attached, byscrew means or similar, to the end of the cylinder 1. Within eachcylinder head 7 a is located an exhaust poppet valve 7 b, coaxiallyaligned with the axis of the cylinder 1. The exhaust poppet valve 7 b isoperated by an exhaust poppet valve actuator 7 c, which may be anelectrically operated solenoid means or other electrical or mechanicalmeans. Accordingly, when the intake poppet valve 6 c and the exhaustpoppet valve 7 b within the first or second chamber 3, 4, are bothclosed, that chamber is effectively sealed and a working fluid containedtherein may be compressed or allowed to expand.

The exhaust means 7 also includes an exhaust manifold channel 7 dprovided within the cylinder head, into which exhaust gases may flow,under the action of a pressure differential between the adjacent firstor second chamber 3, 4 and the fluid within the exhaust manifold channel7 d when the exhaust poppet valve 7 b is open. The flow of the exhaustgases can be better seen in the arrangement of cylinders illustrated inFIG. 20, which shows the direction of the exhaust gas flow to besubstantially perpendicular to the axis of the cylinder 1.

Ignition means 8, such as a spark plug, are also provided at each end ofthe cylinder 1, the ignition means 8 being located within the cylinderhead 7 a and, preferably, recessed such that there is no obstruction ofthe piston 2 during the normal operating cycle of the engine.

The, preferably, coaxial arrangement of the exhaust poppet valve 7 bwith the axis of the cylinder 1 allows the exhaust poppet valve 7 bdiameter to be much larger relative to the diameter of the chambers 3, 4than in a conventional internal combustion engine.

Each cylinder head 7 a is constructed from a hard-wearing and goodinsulating material, such as ceramic, to minimise heat rejection andavoid the need for separate valve seat components.

FIG. 9 shows a time-displacement plot of an engine according to thepresent invention, illustrating the movement of the piston 2 over thecourse of a complete engine cycle. Although the operation of the engineis described here with reference to the first chamber 3, a skilledperson will recognise that the operation and sequence of events of thesecond chamber 4 is exactly the same as the first chamber 3, but 180degrees out of phase—that is to say, top dead centre for the firstchamber 3 occurs at the same time as bottom dead centre for the secondchamber 4.

FIG. 9 a is a table showing a number of different compression ratiocontrol means that may be employed to control the compression ratio inresponse to changes in signals received from a number of differentvariables which can affect the compression ratio during an engine cycle.FIG. 9 b is a flow chart corresponding to FIG. 9 a. The compressionratio control means may comprise part of the control module 9 d,discussed earlier.

Both the table and flow chart illustrate the main variables which canaffect the compression ratio at the different stages (A to F) of anengine cycle, such as the one illustrated in FIG. 9. These variablesinclude: power demand from user, the fuel type being used, thecompression ratio and knock status from the previous engine cycle,piston position, and the kinetic energy of a piston. The table and flowchart illustrate the different processes that take place to control thecompression ratio and how the different variables affect thesethroughout an engine cycle and also the subsequent effect of eachprocess, which can have an effect on more than one of the controlprocesses throughout the engine cycle. It can be seen that in the laststep of the sequence, once the expected compression ratio has beendetermined, optimum ignition timing is achieved by the spark ignitiontiming control means adjusting the timing of the spark event.

The events A to F, highlighted throughout the engine cycle, correspondto the events A to F illustrated in FIG. 10, which shows a typicalpressure-volume plot for a combustion chamber 3, 4 over the course ofthe same engine cycle. The events featured in FIGS. 9 to 10 are referredto in the following discussion of FIGS. 11 to 19.

Considering now a complete engine cycle, at the start of the enginecycle, the first chamber 3 contains a compressed mixture composedprimarily of pre-mixed fuel and air, with a minority proportion ofresidual exhaust gases retained from the previous cycle. It is wellknown that the presence of a controlled quantity of exhaust gases isadvantageous for the efficient operation of the engine, since this canreduce or eliminate the need for intake charge throttling as a means ofengine power modulation, which is a significant source of losses inconventional spark ignition engines. In addition, formation of nitrousoxide pollutant gases are reduced since peak combustion temperatures andpressures are lower than in an engine without exhaust gas retention.This is a consequence of the exhaust gas fraction not contributing tothe combustion reaction, and due to the high heat capacity of carbondioxide and water in the retained gases.

FIG. 11 shows the position of the piston relative to the cylinder 1,defining the geometry of the first chamber 3 at top dead centre (A).This is also around the point of initiation of the combustion phase AB.The distance between the top of the piston 2 b and the end of the firstchamber 3 is at least half the diameter of the first chamber 3, giving alower surface area to volume ratio compared to combustion chambers inconventional internal combustion engines, and reducing the heat lossesfrom the first chamber 3 during combustion. The ignition means 8 arerecessed within the cylinder head 7 a so that in the event that thepiston 2 approaches top dead centre in an uncontrolled manner there isno possibility of contact between the ignition means 8 and the pistoncrown 2 d. Instead, compression will continue until the motion of thepiston 2 is arrested by the continuing build up of pressure due toapproximately adiabatic compression in the first chamber 3. Withreference to FIG. 10, the combustion expansion phase AB is initiated byan ignition event (A).

FIG. 12 shows the position of the piston 2 relative the linear generatormeans 9 mid-way through the expansion phase (AB and BC). The firstchamber 3 expands as the piston 2 moves under the action of the pressuredifferential between the first chamber 3 and the second chamber 4. Thepressure in the second chamber 4 at this point is approximatelyequivalent to the pressure in the intake manifold 6 f. The expansion ofthe first chamber 3 is opposed by the action of the linear generatormeans 9, which may be modulated in order to achieve a desired expansionrate, to meet the engine performance, efficiency and emissionsobjectives.

FIG. 13 shows the position of the piston 2 at bottom dead centrerelative to the first chamber 3. At the end of the expansion phase (C),the motion of the piston 2 is arrested under the action of the lineargenerator means 9 and the pressure differential between the firstchamber 3 and the second chamber 4. The pressure in the second chamber 4at this point is approximately equal to the high pressure in the firstchamber 3 at its top dead centre position (A). Preferably, the expansionratio is at least two times the compression ratio, wherein thecompression ratio is in the range of 10:1 to 16:1. This gives animproved thermal efficiency compared to conventional internal combustionengines wherein the expansion ratio is similar to the compression ratio.

FIG. 14 shows the arrangement of the piston 2 and intake means 6 and theinitial flow of intake gas at the time of bottom dead centre during theintake equalisation phase (CD). This arrangement can also be seen inFIG. 7. At this point, the sliding port intake valve 6 a is open due tothe piston 2 sliding through and past the apertures 1 a, 1 b providedalong the inner wall 1 c of the cylinder 1. The pressure in the firstchamber 3 is lower than the pressure in the intake manifold 6 f due tothe over-expansion reducing fluid pressure in the first chamber 3 anddue to the intake compressor 6 e elevating the pressure in the intakemanifold 6 e. Around this time, the intake poppet valve 6 c is opened byintake poppet valve actuator 6 d allowing intake charge to enter thefirst chamber 3 within cylinder 1 whose pressure approaches equalisationwith the pressure at the intake manifold 6 f. A short time after theintake poppet valve 6 c opens, the exhaust poppet valve 7 b is alsoopened allowing exhaust gases to exit the first chamber 3 under theaction of the pressure differential between the first chamber 3 and theexhaust manifold channel 7 d, which remains close to ambient atmosphericpressure.

FIG. 15 shows the position of the piston 2 during the intake chargedisplacement scavenging phase (DE). Exhaust gas scavenging is achievedby the continuing displacement of exhaust gas in the first chamber 3into the exhaust manifold channel 7 d with fresh intake chargeintroduced at the piston end of the first chamber 3. Once the intendedquantity of intake charge has been admitted to the first chamber 3, theintake poppet valve 6 c is closed and the expulsion of exhaust gascontinues by the movement of the piston 2, as shown in FIG. 17,explained below.

FIG. 16 shows the arrangement of the piston 2 and intake means 6 at thepoint of fuel injection (E). Fuel 5 a is introduced directly onto theapproaching piston crown 2 d which has the effects of rapidly vaporisingfuel, cooling the piston crown 2 d and minimising the losses andemissions of unburned fuel as a wet film on the inner wall 1 c of thecylinder 1, which might otherwise vaporise in the second chamber 4during the expansion phase.

FIG. 17 shows the position of the piston 2 during lubrication (E),whereby a small quantity of lubricant is periodically introduced by thelubrication means 10 directly to the piston outer surface 2 a as itpasses the intake sliding port valve 6 a. This arrangement minimiseshydrocarbon emissions associated with lubricant wetting of the cylinderinner wall, and may also reduce the extent of dissolution of fuel in thecylinder inner wall oil film. Oil control ring features 2 e are includedin the piston crown 2 d and/or bearing elements 2 i to further reducethe extent of lubricant wall wetting in the first and second chambers 3,4.

FIG. 18 shows the position of the piston 2 during the pistondisplacement scavenging phase EF. The intake poppet valve 6 c is closedand the expulsion of exhaust gas continues by the movement of the piston2. The piston 2 at this time is moving towards the exhaust means 7 andreducing the volume of the first chamber 3 due to the combustion eventin the second chamber 4.

As a result of the relatively larger diameter of the exhaust poppetvalve, as discussed above, the limiting area in the exhaust flow pastthe valve stem may approach 40% of the cylinder bore section area,resulting in low exhaust back pressure losses during both the intakecharge displacement scavenging phase (DE) and piston displacementscavenging phase (EF).

FIG. 19 shows a longitudinal section of the position of the piston 2relative to the cylinder 1 mid-way through the compression phase (FA).When a sufficient exhaust gas expulsion has been achieved, such that theproportion of exhaust gas in the fluid in the first chamber 3 is closeto the intended level, the exhaust poppet valve 7 b is closed and thecompression phase (FA) begins. Compression continues at a varying rateas the piston 2 a ccelerates and decelerates under the action of thepressure differential between the first chamber 3 and the second chamber4. The pressure in the second chamber 4 is at this point falling duringthe expansion phases (AB and BC) and by the action of the lineargenerator means 9. The linear generator force may be modulated in orderto achieve the desired compression rate to meet the engine performance,efficiency and emissions objectives. The compression rate in the firstchamber 3 is substantially equal to and opposite the expansion rate inchamber 4.

FIGS. 20A and 20B, in particular, show how the cylinder 1 is,preferably, located coaxially within a cylinder housing 11, whichprovides structural support to the cylinder 1 and can also be arrangedto provide cooling means. The cylinder housing 11 may be slightlyshorter than the cylinder 1 and the cylinder heads 7 a may be attached,by screw fixings or any other suitable means, to the cylinder housing 11to maintain compression between each cylinder head 7 a and the surfaceof each cylinder end 1 d. FIG. 20C shows section of the cylinder housing11 having an electrical machine 9 e attached

The cylinder housing 11 is, preferably, formed by extrusion of a ductilematerial, such as aluminium alloy, and arranged to provide structuralsupport and cooling means 12 whilst allowing the electrical powergenerating components 9 a-9 e to be integrated in close and accuratelydefined proximity to the moving piston 1 within the cylinder 1.

The wall of the cylinder housing 11 extrusion is, ideally, sufficientlythick and/or strong that it is load-bearing to allow a much thinnercylinder 1 wall that provides wear and sealing surfaces than wouldotherwise be required. As mentioned above, the generator of the presentinvention comprises a plurality of magnetic coils 9 a arranged in thecylinder housing 11, a plurality of stators, in the form of magnetisableelements 9 c and the piston 2, which acts as the translator in thisinstance.

The cylinder housing 11, preferably, has sections of material removedalong the length of the cylinder 1 to form one or more recesses 15 that,ideally, extend through the cylinder housing 11 to expose the wall ofthe cylinder 1 housed within. A plurality of the, ideally, load-bearing,magnetisable elements 9 c can be positioned in close proximity to thecylinder 1 by arranging them in the one or more recesses 15, eachmagnetisable element 9 c preferably being fixed directly to the wall ofthe cylinder 1, which separates them from the moving magnetic circuitelements 2 f of the piston 2.

In the example shown, only one magnetisable element 9 c is provided to arecess. However, it should be noted that two or more magnetisableelements 9 c recesses may be positioned within a single recess 15 ifdesired, depending on desired performance characteristics, and that notall recesses 15 have to contain the same number of magnetisable elements9 c, if any. The inner and outer surfaces of the cylinder 1 providesubstrates for wear and thermal coatings respectively. A thermal coatingcan be applied to the outer surface of the cylinder 2 in the form of anadhesive material, for example, to provide a secure, insulating and loadbearing bond between the cylinder 1 and cylinder housing 11.

FIG. 21 shows an exemplary engine arrangement comprising fourfree-piston engines configured to operate in cycles that aresynchronised to create a fully balanced engine. In this configuration,the overall length of the engine generating 50 kw with a thermalefficiency of around 50% is approximately 1400 mm. The cylinder housing11 can be attached, by screw fixings or any other suitable means, to astructural housing 13 which provides the basis for mechanical attachmentof the engine to a vehicle or other device drawing electrical power fromthe electrical output means 9 e such as is shown in FIG. 22. Anenclosure 14 provides a physical enclosure for the engine, manifolds andcontrol systems. Interfaces are provided across the enclosure 14 forintake and exhaust flows, admission of fuel and lubricant, rejection ofheat, output of electrical power and input of electrical power forstart-up and control.

FIG. 23 shows an end view of an arrangement in which a cylinder head 7 ahouses four engines, whereby exhaust gases exit an engine's combustionchamber 3, 4 via the exhaust poppet valve 7 b and flow substantiallyperpendicular to the axes of the cylinders 1.

Advantageously, with the present invention, the narrow bore geometry ofthe first chamber 3, and the relative positions of the intake means 6and exhaust means 7, which are located at opposite ends of the firstchamber 3, permits a highly efficient and effective scavenging processwith little mixing between the intake charge and the exhaust gases. Thisscheme offers several advantages compared to scavenging in conventionaltwo stroke engines or in free piston two stroke engines.

Firstly, the expulsion of exhaust gases can be accurately controlled bythe timing of the exhaust valve closure, providing variable internalexhaust gas recirculation as a means of engine power control without theneed for a throttling device and the associated engine pumping losses.

Secondly, the limited mixing between the retained exhaust gas and theintake charge may improve the completeness of combustion since thecombustion flame front within the fresh charge is not interrupted bypockets of non-combustible exhaust gas mixed with the combustiblefuel/air mixture.

Thirdly, the introduction of fuel 5 a by the fuel injector means 5shortly before the closure of the sliding intake port valve 6 a, andalso the introduction of lubricant by the lubrication means 10 aroundthis time, is unlikely to result in fuel or lubricant entrainment in theexhaust gases and cause tailpipe hydrocarbon emissions.

Furthermore, the geometry of the chambers 3, 4 is such that at top deadcentre, the distance between the top of the piston 2 b and the end ofthe chambers 3, 4 is at least half the diameter of the chamber 3, 4. Therate of change of compression ratio with piston displacement at top deadcentre is therefore smaller than a conventional free piston engine ofsimilar diameter, but in which the depth of the chamber 3, 4 is less. Asa result, the impact of small variations in the depth of the firstchamber 3 at top dead centre due to combustion variations in the secondchamber 4, control system tolerances or other sources of variability,are considerably reduced. Engine operating cycle stability and controlare considerably improved by this feature.

By arresting the motion of the piston 2 at top dead centre (A), adesired compression ratio may be achieved. A target compression ratiomay be in the range 10:1 to 16:1, and higher compression ratios will ingeneral enable higher thermal efficiencies to be achieved. Differentcompression ratio targets may be set for different fuels, to takeadvantage of the octane number characteristics of the particular fuel orblend of fuels in use. Any combination of feedback signals from aknock-sensor, from piston motion, from exhaust gas composition, and fromother engine operating characteristics may be used as input to thecontrol module 9 d in order to achieve the desired compression rate andratio.

An additional benefit of this embodiment compared to other internalcombustion engines is that noise levels are reduced due to theover-expansion cycle and which results in a low pressure differentialacross the exhaust valve immediately prior to opening. As a result, theshock waves propagating through the exhaust system and causing exhaustnoise in a conventional internal combustion engine or free piston engineare substantially avoided.

If the present invention was incorporated into a low cost passengervehicle having a series hybrid drive train configuration, the cost tothe vehicle user as a means for automotive electrical power generationare reduced compared to existing internal combustion engine designs.This reduction in cost is a result of a number of factors, including thelow cost of fuel per unit of electrical power generated due to highthermal efficiency. Other factors include the low cost of componentmanufacture due to the relatively small number of high tolerancedimensions required and hence the low cost of component assembly. Also,the cost of maintenance is low due to the small number of separatecomponents and moving parts required.

Furthermore, the avoidance of complex auxiliary systems and theelimination of complex force transmission pathways including highlystresses hydrodynamic plain bearings characteristic of conventionalinternal combustion engines and the low cost of materials for theengine, due to the reduced part count and the small number of componentshaving functional design constraints that require the use of high costmaterials such as permanent magnets or specialised alloys of aluminiumor steel are all factors that help to keep the cost down.

The thermal efficiency is also improved compared to existing internalcombustion engine designs. In addition to the factors already discussed,the improved efficiency is also a result of good heat exchange,transferring a proportion of the exhaust, engine and electricalgenerator heat losses into the intake charge, reduced frictional lossesdue to the elimination of cylinder wall loads during conversion ofcylinder pressure load to crankshaft torque and the elimination ofthrottling losses due to engine power modulation being achieved byvariable intake charge flow duration at full intake boost pressure andvariable internal exhaust gas recirculation, and not by throttlingintake air flow as is done in a conventional spark ignition engine.

In addition, tailpipe emissions (including NOx, hydrocarbon andparticulate emissions) are reduced compared to other known free pistonengine designs. This reduction in tailpipe emissions is a result of anumber of factors, including: improved control of compression ratio ineach cycle due to the elongated electrical generator geometry, whichresults in a high electrical control authority over piston movementduring the compression stroke and therefore a lower piston displacementerror at top dead centre; and variable retained exhaust gas compositionof compressed charge to reduce peak combustion temperatures andpressures which determine NOx formation.

1.-30. (canceled)
 31. A method of manufacturing an engine, comprisingproviding a cylinder configured to accommodate at least one piston thatis free to reciprocate within the cylinder, extruding a cylinder housingthat is arranged to retain and provide structural support for thecylinder, and securing the cylinder within the cylinder housing suchthat the cylinder wall is reinforced by the structure of the cylinderhousing.
 32. The method of claim 31, further comprising arranging theone or more magnetisable elements to provide load-bearing support to thecylinder.
 33. The method of claim 31, further comprising securing thecylinder within the cylinder housing an adhesive material on the outsideof the cylinder, wherein the adhesive material provides thermalinsulation between the cylinder and cylinder housing.
 34. The method ofclaim 31, further comprising providing the cylinder housing with coolingmeans for cooling the cylinder.
 35. The method of claim 31, furthercomprising coating the interior wall of the cylinder withfriction-reducing material between the interior wall and a pistonpassing along it.
 36. The method of claim 31, further comprisingreducing the thickness of the cylinder wall to a thickness that is lessthan 5% of the cylinder's internal diameter.
 37. The method of claim 36,further comprising limiting the thickness of the cylinder wall to lessthan about 2 mm.
 38. The method of claim 31, further comprisingconstructing the piston using alternating magnetisable elements andnon-magnetisable spacer elements.
 39. The method of claim 31, furthercomprising providing one or more recesses in the cylinder housing thatpermit a plurality of magnetisable elements to be positioned adjacentthe cylinder.